U.S. patent number 6,159,126 [Application Number 09/327,483] was granted by the patent office on 2000-12-12 for toroidal continuously variable transmission.
This patent grant is currently assigned to Nissan Motor Co., Ltd.. Invention is credited to Toshikazu Oshidari.
United States Patent |
6,159,126 |
Oshidari |
December 12, 2000 |
Toroidal continuously variable transmission
Abstract
When a vehicle comprising a toroidal continuously variable
transmission is pulled or run under its own inertia when the engine
has stopped, the rotation torque of the drive wheels varies a
gyration angle of a power roller (44FR, 44FL, 44RR, 44RL) via an
output disk (18, 20) in a direction so as to reduce a speed ratio.
This decrease of speed ratio reduces starting performance when the
vehicle is restarted by the engine. This invention suppresses the
variation of gyration angle by a spring (100) which limits the
displacement of a trunnion (46FR, 46FL, 46RR, 46RL) supporting the
power roller (44FR, 44FL, 44RR, 44RL).
Inventors: |
Oshidari; Toshikazu (Yokosuka,
JP) |
Assignee: |
Nissan Motor Co., Ltd.
(Yokohama, JP)
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Family
ID: |
15987251 |
Appl.
No.: |
09/327,483 |
Filed: |
June 8, 1999 |
Foreign Application Priority Data
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Jun 22, 1998 [JP] |
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10-174932 |
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Current U.S.
Class: |
476/10;
476/42 |
Current CPC
Class: |
F16H
15/38 (20130101) |
Current International
Class: |
F16H
15/32 (20060101); F16H 15/38 (20060101); F16H
015/38 () |
Field of
Search: |
;476/10,42,40 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
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28 47 919 |
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Nov 1978 |
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DE |
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63-92859 |
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Jun 1988 |
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JP |
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Other References
Patent Abstracts of Japan, vol. 011, No. 234 (1987) & JP 62
046060 A (NIPPON SEIKO KK). .
Patent Abstracts of Japan, vol. 011, No. 221 (1987) & JP 62
037562 A (NIPPON SEIKO KK). .
Soviet Inventors Illustrated Section PQ, Week 8711, Mar. 25, 1987
Derwent Publication Ltd., London, GB; Class Q64, AN 87-078306
XP002116873 & SU 1 245 785 A (Svetozakov YU V), Jul. 23,
1987..
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Primary Examiner: Bonck; Rodney H
Attorney, Agent or Firm: Foley & Lardner
Claims
What is claimed is:
1. A toroidal continuously variable transmission, comprising:
an input disk having a rotation axis;
an output disk having the same rotation axis as said input
disk;
a power roller in contact with said input disk and output disk for
transmitting a rotational torque between the disks;
a trunnion for supporting said power roller, said trunnion having a
trunnion axis perpendicular to said rotation axis;
an oil pressure drive device for varying a contact point between
said power roller and said input disk and output disk by displacing
said trunnion within a predetermined range along the trunnion axis;
and
a limiting member for limiting the displacement of said trunnion in
the direction of said trunnion axis, when a rotation torque is
input from said output disk to said power roller,
wherein said limiting member produces an elastic resistance against
a direction of the displacement of said trunnion when there is no
oil pressure from the oil pressure drive device and the rotation
torque is input from said output disk to said power roller.
2. A toroidal continuously variable transmission as defined in
claim 1, wherein said oil pressure drive device comprises a piston
joined to said trunnion and an oil chamber which exerts a pressure
on said piston in the direction of said trunnion axis, and said
limiting member is disposed in said oil chamber.
3. A toroidal continuously variable transmission as defined in
claim 2, wherein said limiting member comprises a plate spring
disposed in said oil chamber for elastically supporting said piston
in a direction opposing said load.
4. A toroidal continuously variable transmission as defined in
claim 3, wherein said trunnion comprises a trunnion shaft coaxial
with said trunnion axis, said piston is fixed to said trunnion
shaft, said trunnion shaft passes through said oil chamber, said
oil chamber has a wall surface facing said piston, and said plate
spring comprises an outer circumferential part which comes in
contact with said wall surface and an inner circumferential part
which comes in contact with said piston.
5. A toroidal continuously variable transmission as defined in
claim 2, wherein said limiting member comprises a ring-shaped wave
spring disposed in said first oil chamber for elastically
supporting said piston in a direction opposing said load.
6. A toroidal continuously variable transmission as defined in
claim 1, wherein said transmission further comprises a resilient
member and said trunnion comprises a second trunnion pushed in the
direction of said trunnion axis by said resilient member, and said
limiting member comprises a supporting member which supports said
second trunnion against said resilient member.
7. A toroidal continuously variable transmission as defined in
claim 6, wherein said oil pressure drive device comprises a piston
connected to said second trunnion and an oil chamber which exerts a
pressure on said piston in the direction of said trunnion axis, and
said supporting member elastically supports said piston in said oil
chamber.
8. A toroidal continuously variable transmission as defined in
claim 7, wherein said second trunnion comprises a trunnion shaft
coaxial with said trunnion axis, and said resilient member
comprises a member which follows the movement of said trunnion
shaft so as to feedback the displacement of said second trunnion to
said oil pressure drive device.
Description
FIELD OF THE INVENTION
This invention relates to a toroidal continuously variable
transmission for a vehicle.
BACKGROUND OF THE INVENTION
Jikkai Sho 63-92859 published by the Japanese Patent Office in 1988
discloses a toroidal continuously variable transmission for a
vehicle.
This toroidal continuously variable transmission comprises power
rollers which transmit a rotation in contact with an input disk and
an output disk, and the rotation speed ratio of the input disk and
output disk is varied by varying the contact points between the
power rollers and the disks. The contact points between a power
roller and the disks are varied by varying a gyration angle of the
power roller by displacing a trunnion which supports the power
roller. The trunnion is displaced according to an oil pressure
supplied via a pressure control valve.
The input disk is connected to an output shaft of an engine via a
forward/reverse change-over mechanism and torque converter. The
output disk is joined to drive wheels via an output gear unit and a
differential gear unit.
SUMMARY OF THE INVENTION
If a vehicle is pulled when an engine has stopped, a rotational
torque is input to an output disk from drive wheels, and this
rotates an input disk and the engine via power rollers.
In this case, the input direction of rotational torque to the
transmission is the reverse of that during normal running. The
rotation resistance of the input disk then varies a gyration angle
of the power roller in the decreasing direction of speed ratio. The
speed ratio mentioned here is equivalent to a value obtained by
dividing the rotational speed of the input disk by the rotation
speed of the output disk. Therefore, as the rotation speed of the
drive wheels relative to engine rotation speed increases the lower
the speed ratio.
On the other hand, during normal start of the vehicle, the speed
ratio is maintained at a maximum value. Due to this, to start the
engine when the vehicle has stopped at a small speed ratio after
being pulled, the speed ratio must be increased to the maximum
value from a small value, and this operation interferes with the
smooth departure of the vehicle. This is the same when the vehicle
is pulled backwards.
It is therefore an object of this invention to limit a variation of
speed ratio in a decreasing direction when a rotational torque is
input into a continuously variable transmission from the output
disk.
It is another object of this invention to eliminate the effect of
other factors causing decrease of speed ratio which is unintended
by a driver.
In order to achieve the above objects, this invention provides a
toroidal continuously variable transmission, comprising an input
disk having a rotation axis, an output disk having the same
rotation axis as the input disk, a power roller in contact with the
input disk and output disk for transmitting a rotational torque
between the disks, a trunnion for supporting the power roller, this
trunnion having a trunnion axis perpendicular to the rotation axis,
an oil pressure drive device for varying a contact point between
the power roller and the input disk and output disk by displacing
the trunnion within a predetermined range along the trunnion axis,
and a limiting member for limiting the displacement of the trunnion
when a load acts on the trunnion in the direction of the trunnion
axis, to a range smaller than the predetermined range when a
rotation torque is input from the output disk to the power
roller.
The details as well as other features and advantages of this
invention are set forth in the remainder of the specification and
are shown in the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a longitudinal sectional view of a vehicle drive unit
comprising a toroidal continuously variable transmission according
to this invention.
FIG. 2 is a cross-sectional view of a first toroidal unit according
to this invention taken along a line II--II of FIG. 1.
FIG. 3 is a cross-sectional view of a second toroidal unit
according to this invention taken along a line III--III of FIG.
1.
FIG. 4 is a longitudinal sectional view of a trunnion according to
another embodiment of this invention.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
Referring to FIG. 1 of the drawings, a vehicle toroidal
continuously variable transmission comprises first and second
toroidal units 10, 11 disposed in series in a transmission case
1.
The rotation of an engine is transmitted to a cam flange 14 via a
torque converter 4, oil pump 5 and forward/reverse change-over
mechanism 9.
The rotation of the cam flange 14 is transmitted to an input disk
17 of the first toroidal unit 10 via a cam rollers 15.
The input disk 17 is joined to an input disk 19 of the second
toroidal unit 11 via a rotation shaft 3. These input disks 17, 19
are joined to the rotation shaft 3 via ball splines 16, 21.
Rotation relative to the shaft 3 is restricted, and axial
displacement is permitted within a predetermined range.
The cam rollers 15 exert a thrust load according to the rotation of
the cam flange 14, on the input disks 17, 19, and the input disks
17, 19 are pushed towards output disks 18, 20 facing the input
disks 17, 19. The output disks 18, 20 are engaged free to rotate on
the outer circumference of the rotation shaft 3.
The input disk 17 and output disk 18 forming the first toroidal
unit 10 have corresponding toroidal-shaped wall surfaces 17A, 18A,
and a pair of power rollers 44FR, 44FL are gripped by the wall
surfaces 17A, 18A due to the aforesaid thrust load. Identical power
rollers 44RR, 44RL are gripped between the input disk 19 and output
disk 20 of the second toroidal unit 11, as shown in FIG. 3.
The reason why the symbol L is given to the right power roller and
the symbol R is given to the left power roller in FIGS. 2 and 3 is
that the symbols R are given to the parts situated on the right
hand and the symbols L are given to the parts situated on the left
hand when they are viewed from the right hand of FIG. 1. FIGS. 2
and 3 are both cross-sectional views viewed from the left hand of
FIG. 1, so the symbols R are given to the left parts and the
symbols L are given to the right parts.
The rotation of the input disks 17, 19 is transmitted to the output
disks 18, 20 via these power rollers 44FR, 44FL, 44RR and 44RL.
The rotation of the output disks 18, 20 is transmitted to an output
shaft and drive wheels, not shown, via an output gear 22, gear 25,
counter shaft 27 and gear 28.
The contact point of the input disk 17 and output disk 18 of the
power roller 44FR (44FL) varies according to the gyration angle of
the power roller 44FR (44FL) i.e., the rotation angle of the power
roller 44FR (44FL) about an axis O.sub.3 in FIG. 2, and the ratio
of rotation speeds of the input disk 17 and output disk 18 is
determined according to the distance between the contact point and
an axis O.sub.1. This speed ratio is the speed ratio of this
toroidal transmission. The same relation holds in the power rollers
44RR (44RL), input disk 19 and output disk 20 of the second
toroidal unit 11.
Next, the construction of the parts of the toroidal unit will be
described referring to FIGS. 2 and 3.
The first toroidal unit 10 shown in FIG. 2 comprises a pair of
trunnions 46FR, 46FL for supporting the power rollers 44FR, 44FL.
The power rollers 44FR, 44FL are supported via crank-shaped
eccentric shafts 29 by the trunnions 46FR, 46FL. The eccentric
shaft 29 comprises a base end supported free to rotate by a
trunnion 46 and a tip end which is crank-shaped. The power roller
44FR (44FL) is supported free to rotate around a rotation shaft
O.sub.2 by this point. The power roller 44FR (44FL) is also
permitted to swing within predetermined limits around the base end
as fulcrum.
The upper parts of the trunnions 46FR, 46FL are connected by an
upper link 50 via spherical joints, and the lower parts of the
trunnions 46FR, 46FL are connected by a lower link 52 via spherical
joints.
A hole is formed in the lower part of the trunnion 46FR(46FL), and
a trunnion shaft 70 is inserted into this hole and 46FR(46FL)
joined to the trunnion 46FR (46FL) by a pin 56.
Aboss 78A of a servo piston 78FR(78FL) engages with the outer
circumference of the trunnion shaft 70, and the servo piston
78FR(78FL) is secured with the trunnion 46FL by tightening a nut 82
which screws onto a male screw formed on the lower end of the
trunnion shaft 70.
The servo piston 78FR is housed in a piston housing 60. In the
piston housing 60, a first oil chamber 92A is formed above the
servo piston 78FR, and a second oil chamber 92B is formed below it.
The servo piston 78FL is also housed in the same piston housing 60,
but unlike the case of the servo piston 78FR, a first oil chamber
90A is formed below and a second oil chamber 90B is formed above
the servo piston 78FL.
Equal oil pressures are supplied to the first oil chambers 90A, 92A
via an oil pressure control valve, not shown. Equal oil pressures
are also supplied to the second oil chambers 90B, 92B, from the
same oil pressure control valve. The servo pistons 78FL, 78FR are
displaced along the axis O.sub.3 according to the differential
pressure of the first oil chambers 90A, 92A and the second oil
chambers 90B, 92B. The servo pistons 78FL, 78FR therefore displace
in opposite directions to each other.
Due to this displacement, each of the trunnions 46FR, 46FL suffers
a rotational displacement around the axis O.sub.3 together with the
power roller 44FR (44FL). The rotational displacements of the power
rollers 44FR and 44FL take place in reverse directions.
According to the change in the distance of the contacts from the
axis O.sub.1, the contact points of the input disk 17 and the
output disk 19 with the power rollers 44FR, 44FL, change, and the
ratio of speeds of the input disk 17 and output disk 19
changes.
In the trunnion 46FR, a precess cam 66 is further gripped by the
boss 78A and the nut 82.
The precess cam 66 comprises a slanting guide groove 66A on its
outer circumference, and one end of a link 68 is inserted in the
guide groove 66A. Consequently, the displacement of the trunnion
46FR in the direction of the axis O.sub.3, and the rotational
displacement, i.e., the gyration angle, of the power roller 44FR
around the axis O.sub.3, are fed back to an oil pressure control
valve via the displacement of the link 68. The link 68 is pushed in
the direction of the arrow Fa in the figure by a spring, not shown,
so that an upper surface of the end of the link 68 and a guide
surface of the guide groove 66A are always in contact.
In the normal running state of the vehicle when an engine of the
vehicle is driving the drive wheels via the continuously variable
transmission, if the pressures of the first oil chambers 90A, 92A
are raised and the pressures of the second oil chambers 90B, 92B
are reduced, the servo pistons 78FR, 78FL displace in the shift-up
direction shown by the solid arrow S.sub.U in the Figure along the
axis O.sub.3 due to the pressure differential.
As a result, the gyration angles of the power rollers 44FR, 44FL
change, and the speed ratio varies in the decreasing direction,
i.e., towards higher speed.
Conversely, if the pressures of the first oil chambers 90A, 92A are
reduced and the pressures of the second oil chambers 90B, 92B are
increased, the servo pistons 78FR, 78FL displace in the shift-down
direction shown by the broken arrow S.sub.D in the figure. As a
result, the gyration angles of the power rollers 44FR, 44FL change,
and the speed ratio varies in the increasing direction, i.e.,
towards lower speed.
The construction of the second toroidal unit 11 shown in FIG. 3 is
almost the same as that of the first toroidal unit 10. However, in
the second toroidal unit 11, there is no feedback mechanism
comprising a precess cam and link.
An oil pressure equal to that supplied to the oil chambers 90A, 92A
of the first toroidal unit 10, is supplied to the oil chambers 94A,
96A of the second toroidal unit 11, and an oil pressure equal to
that supplied to the oil chambers 90B, 92B of the first toroidal
unit 10 is also supplied to the oil chambers 94B, 96B. Due to this,
the first toroidal unit 10 and second toroidal unit 11 transmit the
rotation of the input disks 17, 19 to the output disks 18, 20 at
the same speed ratio.
In this continuously variable transmission, first plate springs 100
are respectively arranged in the second oil chamber 90B of the
first toroidal unit 10 and the second oil chambers 94B and 96B of
the second toroidal unit.
A second plate spring 102 is arranged in the first oil chamber 92A
of the first toroidal unit 10.
The plate springs 100, 102 are both fitted to the outer
circumference of the boss 78A. The inner circumferential diameters
of the plate springs 100, 102 are set so as to leave a small
clearance between the inner circumference of the plate springs 100,
102 and the boss 78A. The inner circumference of the plate springs
100, 102 comes in contact with the servo pistons 78FL, 78RR and
78RL. The rims of the plate springs 100, 102 come in contact with
the wall surface 60A of the piston housing 60 which forms the
second oil chambers 90B, 94B and 96B. Since a part of large
diameter comes in contact with the wall surface 60A of the case 60,
the unit pressure exerted by the plate springs 100, 102 on the wall
surface 60A can be reduced, and antiwear resistance of parts is
improved. Further, the provision of a slight clearance between the
inner circumference of the plate springs 100, 102 and the boss 78A
has a desirable effect in that it makes the direction of the
elastic restoring force of the plate springs 100, 102 which acts on
the servo pistons 78FR, 78FL, 78RR, 78RL, coincide with the
direction of the axis O.sub.3.
The first plate spring 100 pushes the servo pistons 78FL, 78RR and
78RL in the downshift direction. Let the sum total of these forces
be Fs.sub.1. The second plate spring 102 pushes the servo piston
78FR in the upshift direction. Let this force be Fs.sub.2.
If the vehicle is pulled forwards when the engine has stopped, a
rotational torque is input from the drive wheels to the output
disks 18, 20 of the continuously variable transmission, and this
rotates the input disks 17, 19 via the power roller 44FR, 44FL,
44RR and 44RL. At this time, the rotation resistance of the input
disks 17, 19 exerts a force in the upshift direction on the power
roller 44FR, 44FL, 44RR and 44RL.
Herein, let the force exerted in the upshift direction by the input
disks 17, 19 on the power rollers 44FR, 44FL, 44RR and 44RL be
F.sub.1. The relation between this force F.sub.1 and the force
Fs.sub.1 in the downshift direction exerted by the plate spring 100
on the servo pistons 78FL, 78RR and 78RL is set as follows.
However, as the force Fs.sub.1 depends on the deformation amount of
the plate spring 100, for example, the spring load of the plate
spring 100 is set so that, for example, the following relation
exists relative to a displacement amount of the servo pistons 78FL,
78RR and 78RL which is equivalent to a speed ratio equal to at
least X:1.
Also, let the force in the downshift direction with which the
spring pushing the link 68 exerts on the servo piston 78FR be Fa.
The spring load of the plate spring 102 is set so that the
following relation holds between this force Fa, the force F.sub.1
in the upshift direction acting on the power roller 44Fr, and the
force Fs.sub.2 which the second plate spring 102 exerts on the
servo piston 78FR.
wherein.apprxeq.means "approximately equal to."
By setting the characteristics of the first plate spring 100 in
this way, when the vehicle is pulled forward, the speed ratio of
the continuously variable transmission does not become less than
the set value X:1, and impairment of startup performance when the
vehicle is restarted, is therefore prevented.
By balancing the force F.sub.1 in the upshift direction acting on
the power roller 44FR and the elastic restoring force Fs.sub.2 of
the second plate spring 102, the load conditions of the power
roller 44FR become identical to those of the other power rollers
44FL, 44RR and 44RL. Accordingly, the operation of the power roller
44FR may be synchronized with those of the other power rollers
44FL, 44RR and 44RL.
The first and second plate springs 100, 102 are housed inside the
piston housing 60, so they do not take up extra space.
Next, another embodiment of this invention will be described
referring to FIG. 4.
According to this embodiment, a wave spring 104 is used instead of
the first plate spring 100.
The wave spring 104 is a ring-shaped spring which has a wave-shaped
cross-section in a circumferential direction. The upward crest of
the wave spring 104 comes in contact with the piston housing 60,
and the lower crest comes in contact with the upper surface of the
servo piston 78FL in the case of trunnions 46FL and 46RR as shown
in FIG. 4. In the trunnion 46RL, the wave spring 104 is disposed in
the lower oil chamber 94B, its upper crest comes in contact with
the lower surface of the piston 78RL and its lower crest comes in
contact with the piston housing 60. As in the case of the first
plate spring 100, the wave spring 104 is attached leaving a small
clearance with the outer circumference of the boss 78A. Guides 104B
are also formed in the wave spring. One of the guides 104B comes in
contact with the piston housing 60 while the other of the guides
104B comes in contact with the servo piston 78FL (78RR, 78RL) so as
to ensure coaxiality of the wave spring 104 with the boss 78A.
In this embodiment, the same effect is obtained as in the aforesaid
first embodiment. Further, the contact area with the case 60 and
the upper surfaces of the servo pistons 8FL, 78RR and 78RL is
larger than when using the first plate spring 100, so wear of parts
is prevented and durability of the transmission is further
improved.
Both of the above-mentioned embodiments aim to limit the change of
speed ratio when the vehicle is pulled in the forward direction.
However, the change of speed ratio when the vehicle is being pulled
or running under its own inertia in the reverse direction can also
be suppressed by reversing the direction of the elastic restoring
force of the first plate spring 100, i.e., by for example disposing
the first plate spring 100 of the first oil chamber 90B, in the
second oil chamber 90A.
The contents of Tokugan Hei 10-174932, with a filing date of Jun.
22, 1998 in Japan, are hereby incorporated by reference.
Although the invention has been described above by reference to
certain embodiments of the invention, the invention is not limited
to the embodiments described above. Modifications and variations of
the embodiments described above will occur to those skilled in the
art, in light of the above teachings.
The embodiments of this invention in which an exclusive property or
privilege is claimed are defined as follows:
* * * * *